U.S. patent application number 12/387536 was filed with the patent office on 2010-11-04 for radial compressor with blades decoupled and tuned at anti-nodes.
This patent application is currently assigned to Hamilton Sundstrand Corporation. Invention is credited to Loc Q. Duong, Shiv C. Gupta, Xiaolan Hu.
Application Number | 20100278633 12/387536 |
Document ID | / |
Family ID | 43030468 |
Filed Date | 2010-11-04 |
United States Patent
Application |
20100278633 |
Kind Code |
A1 |
Duong; Loc Q. ; et
al. |
November 4, 2010 |
Radial compressor with blades decoupled and tuned at anti-nodes
Abstract
A gas turbine engine includes a radial compressor with first and
second blades. The first and second blades have tuned leading edges
that prevent natural frequencies from exciting at speeds within an
expected operating speed range.
Inventors: |
Duong; Loc Q.; (San Diego,
CA) ; Gupta; Shiv C.; (San Diego, CA) ; Hu;
Xiaolan; (San Diego, CA) |
Correspondence
Address: |
KINNEY & LANGE, P.A.
THE KINNEY & LANGE BUILDING, 312 SOUTH THIRD STREET
MINNEAPOLIS
MN
55415-1002
US
|
Assignee: |
Hamilton Sundstrand
Corporation
Rockford
IL
|
Family ID: |
43030468 |
Appl. No.: |
12/387536 |
Filed: |
May 4, 2009 |
Current U.S.
Class: |
415/119 ;
29/889.21 |
Current CPC
Class: |
F04D 29/30 20130101;
Y10S 416/50 20130101; F04D 29/284 20130101; F04D 29/666 20130101;
F05D 2240/303 20130101; Y10T 29/49321 20150115 |
Class at
Publication: |
415/119 ;
29/889.21 |
International
Class: |
F04D 29/66 20060101
F04D029/66; B23P 15/04 20060101 B23P015/04 |
Claims
1. A radial compressor for use in a gas turbine engine operating in
an expected operating speed range, the radial compressor
comprising: a first blade having a first leading edge with a first
normal portion and a first tuned portion, wherein the first tuned
portion has a thickness different than that of the first normal
portion; a second blade having a second leading edge with a second
normal portion and a second tuned portion, wherein the second tuned
portion has a thickness different than that of the second normal
portion; and a disc connecting the first blade to the second blade
and having a thickness sufficient to decouple vibration in the
first blade from vibration in the second blade when operating in
the expected operating speed range.
2. The radial compressor of claim 1, wherein the first blade has a
trailing edge, wherein the disc has a rim at its outer diameter,
and wherein the rim has a thickness greater than about 1.3 times a
thickness of the trailing edge.
3. The radial compressor of claim 1, wherein thicknesses of the
first and second tuned portions are sufficiently different from
thicknesses of the first and second normal portions to tune natural
frequencies of the first and second blades outside of the expected
operating speed range.
4. The radial compressor of claim 1, wherein the first and second
tuned portions cause the first and second blades, respectively, to
have first and second natural frequencies that excite at operating
speeds greater than the expected operating speed range.
5. The radial compressor of claim 1, wherein the first tuned
portion causes the first blade to have a first natural frequency
that excites at a first operating speed below the expected
operating speed range and wherein the second tuned portion causes
the second blade to have a second natural frequency that excites at
a second operating speed greater than the expected operating speed
range.
6. The radial compressor of claim 1, wherein the first blade is one
of a plurality of substantially similar splitter blades and the
second blade is one of a plurality of substantially similar main
blades.
7. The radial compressor of claim 1, wherein the first and second
tuned portions are positioned to prevent formation of first and
second vibration anti-nodes at the first and second tuned portions
at speeds within the expected operating speed range.
8. The radial compressor of claim 1, wherein the first and second
tuned portions are positioned further from the disc than the first
and second normal portions, respectively.
9. The radial compressor of claim 1, wherein the first tuned
portion is thinner than the first normal portion.
10. The radial compressor of claim 9, wherein the second tuned
portion is thinner than the second normal portion.
11. The radial compressor of claim 9, wherein the second tuned
portion is thicker than the second normal portion.
12. The radial compressor of claim 1, wherein the radial compressor
is an impeller for a gas turbine engine.
13. A gas turbine engine comprising: a radial compressor having
first and second blades with tuned leading edges that prevent
natural frequencies from exciting at speeds within an expected
operating speed range.
14. The radial compressor of claim 13, wherein the radial
compressor includes a disc connecting the first blade to the second
blade and having a thickness sufficient to decouple vibration in
the first blade from vibration in the second blade when operating
in the expected operating speed range.
15. A method for tuning a radial compressor, the method comprising:
designing the radial compressor to have a first blade connected to
a second blade by a disc, wherein the first and second blades have
first and second blade resonant modes that excite in an expected
operating speed range of the radial compressor; modifying the disc
to have a stiffness sufficient to reduce transmission of vibration
between the first and second blades when operating in the expected
operating speed range; tuning the first and second blades by
modifying mass quantity at primary anti-nodes of the first and
second blade resonant modes, respectively; and fabricating the
radial compressor as modified and tuned.
16. The method of claim 15, wherein the first blade has a trailing
edge, wherein the disc has a rim at its outer diameter, and wherein
modifying the disc causes the rim to have a thickness greater than
about 1.3 times a thickness of the trailing edge.
17. The method of claim 15, wherein the disc is modified by
increasing thickness of the disc.
18. The method of claim 15, wherein the step of designing the
radial compressor includes creating an electronic model of the
radial compressor.
19. The method of claim 15, wherein the steps of modifying and
tuning occur electronically.
20. The method of claim 15, wherein the primary anti-nodes are
positioned at first and second leading edges of the first and
second blades, respectively.
21. The method of claim 20, wherein the first blade is tuned by
decreasing mass at the primary anti-node on the first blade.
22. The method of claim 20, wherein the second blade is tuned by
decreasing mass at the primary anti-node on the second blade.
23. The method of claim 15, and further comprising: identifying the
primary anti-nodes of the first and second blades through
eigenvalue solutions.
24. The method of claim 15, wherein the primary anti-node on the
first blade has a greater deflection than all other anti-nodes of
the first blade resonant mode and the primary anti-node on the
second blade has a greater deflection than all other anti-nodes of
the second blade resonant mode.
Description
CROSS-REFERENCE TO RELATED APPLICATION
[0001] Reference is made to application Ser. No. ______ entitled
"RADIAL COMPRESSOR OF ASYMMETRIC CYCLIC SECTOR WITH COUPLED BLADES
TUNED AT ANTI-NODES", which is filed on even date and is assigned
to the same assignee as this application.
[0002] Reference is also made to application Ser. No. 11/958,585
entitled "METHOD TO MAXIMIZE RESONANCE-FREE RUNNING RANGE FOR A
TURBINE BLADE", filed on Dec. 18, 2007 by Loc Q. Duong, Ralph E.
Gordon, and Oliver J. Lamicq and is assigned to the same assignee
as this application.
BACKGROUND
[0003] The present invention relates to radial compressors, and in
particular, to radial compressors with blades tuned according to
natural frequency.
[0004] Gas turbine engines typically include several sections such
as a compressor section, a combustor chamber, and a turbine
section. In some gas turbine engines, the compressor section
includes a radial compressor with a series of main blades and
splitter blades connected by a disc. During operation of the gas
turbine engine, the main blades and splitter blades can be subject
to vibratory excitation at frequencies which coincide with integer
multiples, referred to as harmonics, of the radial compressor's
rotational frequency. As a result of the vibratory excitation, the
main blades and/or the splitter blades can undergo vibratory
deflections that create vibratory stress on the blades. If the
vibratory excitation occurs in an expected operating speed range of
the radial compressor, the vibratory stresses can create high cycle
fatigue and cracks over time.
SUMMARY
[0005] According to the present invention, a gas turbine engine
includes a radial compressor with first and second blades. The
first and second blades have tuned leading edges that prevent
natural frequencies from exciting at speeds within an expected
operating speed range.
[0006] Another embodiment includes a method for tuning a radial
compressor. The method includes designing the radial compressor to
have a first blade connected to a second blade by a disc, wherein
the first and second blades have first and second blade resonant
modes that excite in an expected operating speed range of the
radial compressor, modifying the disc to have a stiffness that
reduces transmission of vibration between the first and second
blades, tuning the first and second blades by modifying mass
quantity at primary anti-nodes of the first and second blade
resonant modes, and fabricating the radial compressor as modified
and tuned.
BRIEF DESCRIPTION OF THE DRAWINGS
[0007] FIG. 1 is a perspective view of a radial compressor.
[0008] FIG. 2A is rear view of the radial compressor of FIG. 1,
showing deflection of a resonant mode shape.
[0009] FIG. 2B is a simplified schematic view of the resonant mode
shape of FIG. 2A.
[0010] FIG. 3 is a nodal diameter interference map.
[0011] FIG. 4 is a flow chart of a method of tuning the radial
compressor of FIG. 1.
[0012] FIG. 5 is an enlarged view of a cyclic sector of the radial
compressor of FIG. 1.
[0013] FIG. 6 is a schematic sectional view of an alternative
embodiment of the cyclic sector of the radial compressor taken
along line 66 of FIG. 5.
DETAILED DESCRIPTION
[0014] FIG. 1 is a perspective view of radial compressor 10 (also
called an impeller or a bladed disc). Radial compressor 10 includes
a plurality of blades 12 connected to disc 14 (also called a body).
Disc 14 is curved and substantially frusto-conical, extending from
hub 16 at its inner diameter to rim 18 at its outer diameter.
Blades 12 includes a series of splitter blades (e.g. splitter blade
20) positioned alternately with a series of main blades (e.g. main
blade 22). Splitter blade 20 has a different shape, including a
shorter chord length, than that of main blade 22. Splitter blade 20
and main blade 22 each have fixed edge 24 attached to disc 14 and
free edge 26 unattached. Free edge 26 includes leading edge 28,
trailing edge 30, and side edge 32 there-between.
[0015] Hub 16 can be attached to a compressor shaft of a gas
turbine engine (not shown). In operation, air from a turbine inlet
(not shown) can pass over leading edge 28, is compressed by blades
12 as radial compressor 10 rotates, and passes over trailing edge
30 on its way to a combustion chamber (not shown). Because
operation of gas turbine engines is well known in the art, it will
not be described in detail herein. However, during engine
operation, various aero-excitation source frequencies can be
created as air passes over components of the gas turbine engine,
such as inducer or exducer vanes. Different source frequencies can
be created at different operating speeds. These source frequencies
are transmitted to the air, causing unsteady fluid pressure, and
can then be transmitted to radial compressor 10. Radial compressor
10 can have one or more natural frequencies (also called resonance
frequencies) in which one or more blades 12 and/or disc 14 will
vibrate. If a natural frequency coincides with an aero-excitation
source frequency, an interference can occur, causing undesired
harmonic vibration. A variety of possible blade anti-nodes 34 are
illustrated on free edges 26 of blades 12. Primary anti-node 35 is
that with the greatest deflection of all blade anti-nodes 34 on a
particular blade 12. If a particular blade 12 has two anti-nodes 34
with almost the same deflection, both can be referred to as primary
anti-nodes 35, and any other anti-nodes 34 can be referred to as
secondary anti-nodes 34.
[0016] FIG. 2A is rear view of radial compressor 10, showing
deflection of a resonant mode shape of disc 14. In the illustrated
resonant mode shape, eight disc anti-nodes 36 are present. Disc
anti-nodes 36 are points of greatest deflection of disc 14 in this
resonant mode shape.
[0017] FIG. 2B is a simplified schematic view of the mode shape of
FIG. 2A. Nodal diameters 38A-38D divide disc anti-nodes 36. While
disc anti-nodes 36 (shown in FIG. 2A) are points of greatest
deflection, nodal diameters 38A-38D are lines of approximately zero
deflection during harmonic vibration. The "+" and "-" symbols
illustrate direction of deflection for disc anti-nodes 36 at a
given moment in time. Deflection caused by harmonic vibration of
disc 14 is transmitted to, and combines with deflection of, blades
12 (shown in FIG. 1).
[0018] FIG. 3 illustrates nodal diameter (ND) interference map 50.
ND interference map 50 plots potential interferences associated
with various nodal diameters against vibration frequency. Along the
horizontal axis of ND interference map 50, nodal diameters are
identified as n-1, n, n+1, etc. Along the vertical axis, vibration
frequency is plotted. Upper bound line 52 and lower bound line 54
are upper and lower bounds of an expected operating speed range of
a gas turbine engine. Because gas turbine engines tend to operate
within their expected operating speed ranges, vibration
interferences that occur within the expected operating speed range
can be of particular importance.
[0019] For example, radial compressor 10 has a variety of natural
frequencies associated with nodal diameter n that are potentially
excitable at different operating speeds. However, radial compressor
10 only has two natural frequencies 56 and 58 associated with nodal
diameter n that occur in the expected operating speed range. As
illustrated, natural frequency 56 corresponds to splitter blade 20
and natural frequency 58 corresponds to main blade 22. It can be
desirable to tune radial compressor 10 such that natural
frequencies 56 and 58 excite outside of the expected operating
speed range. For example, radial compressor 10 could be tuned such
that natural frequencies 56' and 58' occur below lower bound line
54. In that case, natural frequencies 56' and 58' will not be
excited in the expected operating speed range. Natural frequencies
56' and 58' could, however, be excited for a period of time as the
gas turbine engine speeds up during initial startup and shutdown.
Alternatively, radial compressor 10 could be tuned such that
natural frequencies 56'' and 58'' occur above upper bound line 52.
In that case, natural frequencies 56'' and 58'' will not be excited
in the expected operating speed range nor during initial startup
and shutdown. In further alternative, radial compressor 10 could be
tuned such that natural frequency 56' occurs below lower bound line
54 and natural frequency 58'' occurs above upper bound line 52.
[0020] FIG. 4 is a flow chart of a method of tuning radial
compressor 10. The method begins by designing a radial compressor,
such as radial compressor 10 of FIG. 1, that requires tuning (step
100). In step 100, radial compressor 10 can be physically
fabricated, or an electronic model of radial compressor 10 can be
created. Next, an expected operating speed range for radial
compressor 10 is determined (step 102). For example, radial
compressor 10 could be expected to operate in a particular gas
turbine engine in a speed range of between about 15,300 revolutions
per minute (RPM) and about 15,900 RPM. Then aero-excitation source
frequencies in the expected operating speed range are determined
(step 104). The aero-excitation source frequencies coincide with
integer multiples of the engine operating speed (the rotational
frequency of radial compressor 10). Next, blade resonant mode
shapes which have interferences are determined (step 106). An
interference occurs when one of blades 12 has a resonant mode with
a natural frequency that coincides with one of the aero-excitation
source frequencies at a particular nodal diameter n. In some
circumstances (such as that illustrated above with respect to FIG.
3), splitter blade 20 and main blade 22 will each have a different
blade resonant mode with a corresponding natural frequency that
coincides with one of the aero-excitation source frequencies within
the expected operating speed range.
[0021] Once the blade resonant modes are identified, stiffness of
disc 14 is modified to reduce transmission of vibration between
blades 12 (step 108). Prior art discs can be relatively thin,
allowing vibration in one blade, such as splitter blade 20, to be
easily transmitted to and excite another nearby blade, such as main
blade 22. This effect couples blade vibrations together such that
modifications to splitter blade 20 also affect natural frequency of
main blade 22. This coupling can make it difficult to predictably
tune a given blade. Thickness of disc 14 can be increased to
stiffen disc 14 in order to reduce transmission of vibration
between splitter blade 20 and main blade 22. For example, thickness
of disc 14 can be increased at rim 18 to a thickness greater than
about 1.3 times a thickness of trailing edge 30 of one of blades
12. If disc 14 is connected to blades 12 with a tapered fillet
portion (not shown) at fixed edge 24, thickness of trailing edge 30
is measured at a normal portion of trailing edge 30, not the
tapered portion. Thickness can be increased until vibrations
between splitter blade 20 and main blade 22 are substantially
decoupled when operating in the expected operating speed range.
After decoupling, vibrations in splitter blade 20 will not excite
resonant vibrations in main blade 22, and vise versa. Decoupling
can be performed using a finite element method.
[0022] After splitter blade 20 and main blade 22 are decoupled,
location of blade anti-nodes 34 of the blade resonant mode shapes
with interferences are identified on each of splitter blade 20 and
main blade 22 (step 110). Blade anti-nodes 34 typically occur along
free edge 26, and in particular, along leading edge 28. If there is
more than one blade anti-node 34 along free edge 26, one or more
primary anti-nodes 35 have greater deflection than all other blade
anti-nodes 34 of the blade resonant mode shape in question. In
radial compressors such as radial compressor 10, one primary
anti-node 35 is typically positioned along leading edge 28.
Location of blade anti-nodes 34 can be determined through
eigenvalue solutions, in a manner known in the art.
[0023] Then splitter blade 20 and main blade 22 are tuned at blade
anti-nodes 34 (step 112). Tuning is performed by modifying mass
localized at one or more blade anti-nodes 34 on each of splitter
blade 20 and main blade 22. Increasing mass at blade anti-node 34
decreases natural frequency, and decreasing mass at blade anti-node
34 increases natural frequency. Mass can be modified until the
natural frequency of the blade resonant mode shapes that have
interferences are moved out of the expected speed range. Mass can
be further modified to further increase a substantially
resonance-free running range at the nodal diameter at issue.
Because splitter blade 20 is vibrationally decoupled from main
blade 22, each blade can be independently tuned without mistuning
the other. Step 112 can be repeated to tune all of blades 12. It
can be relatively effective and efficient to modify mass only at
primary anti-node 35 on each leading edge 28 of blades 12. If
further tuning is desired, mass quantity can be modified on one or
more of blades 12 at an additional blade anti-node. After tuning is
complete, radial compressor 10 can have no natural frequencies that
excite in the expected operating speed range. Leading edges 28 are
tuned to prevent natural frequencies from exciting at speeds within
the expected operating speed range.
[0024] Some or all of steps 100-112 can be performed physically,
electronically, or both. If steps 100-112 are performed
electronically, radial compressor 10 can then be fabricated as
electronically modified and tuned. Radial compressor 10 can be
fabricated using techniques such as forging and machining.
[0025] FIG. 5 is an enlarged sectional view of cyclic sector 200,
which is one of a plurality of duplicate sectors of radial
compressor 10 and has been modified as described with respect to
the method of FIG. 4. Cyclic sector 200 includes splitter blade 20'
and main blade 22' connected by disc 14'. Disc 14' is similar to
disc 14 of FIG. 1 except that disc 14' is sufficiently thick to
decouple vibration between splitter blade 20' and main blade 22'.
Splitter blade 20' is similar to splitter blade 20 of FIG. 1 except
that leading edge 28' of splitter blade 20' has normal portion 202
and tuned portion 204. Main blade 22' is similar to main blade 22
of FIG. 1 except that leading edge 28' of main blade 22' has normal
portion 206 and tuned portion 208. Tuned portions 204 and 208 are
positioned at locations that coincided with anti-nodes prior to
tuning, and prevent formation of those anti-nodes at speeds within
the expected operating speed range. Tuned portions 204 and 208 can
be described as a notch, where mass is trimmed to increase natural
frequencies of blade modes of each of splitter blade 20' and main
blade 22'. In the illustrated embodiment, tuned portions 204 and
208 are positioned radially further from disc 14 than normal
portions 202 and 206.
[0026] FIG. 6 is a schematic sectional view of an alternative
embodiment of cyclic sector 200'' of radial compressor 10 taken
along line 6-6 of FIG. 5. Cyclic sector 200'' of FIG. 6 is similar
to cyclic sector 200 of FIG. 5 except for mass modification at
tuned portions 204'' and 208''. In the illustrated embodiment, mass
removal can be achieved by smoothly and continuously reducing
thickness of each of splitter blade 20'' and main blade 22'' at
tuned portions 204'' and 208''. Tuned portions 204'' and 208'' are
thinner than normal portions 202 and 206, respectively. Non-tuned
thicknesses 210 and 212 (a thickness of tuned portions 204'' and
208'' prior to tuning) are substantially equal to thicknesses of
normal portions 202 and 206, respectively. The locations of tuned
portions 204'' and 208'' would coincide with anti-nodes if tuned
portions 204 and 208 had thicknesses substantially equal to those
of normal portions 202 and 206, respectively.
[0027] Splitter blade 20'' and main blade 22'' can also be modified
by adding mass at tuned portions 204'' and 208''. For example, mass
addition can be achieved by smoothly and continuously increasing
thickness of splitter blade 20'' at tuned portion 204'' from
non-tuned thickness 210 to increased mass tuned thickness 214.
Smooth mass modification allows for reduced aerodynamic impact and
flow separation.
[0028] After splitter blade 20'' and main blade 22'' are tuned,
each blade's contour profile geometry can be optimized to reduce
stress concentration while maintaining a desirable aero-constraint
on an incident angle of leading edge 28'' within about 2 degrees.
All of radial compressor 10 can be tuned similarly to cyclic sector
200'' such that main blade 22'' is one of a plurality of
substantially similar tuned main blades and splitter blade 20'' is
one of a plurality of substantially similar tuned splitter
blades.
[0029] It will be recognized that the present invention provides
numerous benefits and advantages. For example, tuning radial
compressor 10 moves natural frequencies out of an expected
operating speed range and, therefore, reduces vibratory stresses
and cracks in radial compressor 10. By increasing thickness of disc
14, splitter blade 20 and main blade 22 can be decoupled and,
consequently, independently tuned. By modifying mass at primary
anti-nodes 35 on splitter blade 20 and main blade 22, tuning can be
more efficient and more effective than by modifying mass at other
locations on blades 12, disc 14, or elsewhere in the gas turbine
engine. Additionally, by modifying mass at leading edges 28 instead
of at trailing edges 30, problems associated with mass modification
at trailing edge 30 can be reduced (such as weakening the blades
due to elastic deformation if trailing edge 30 is made thinner or
increasing steady state stress if trailing edge 30 is made
thicker).
[0030] While the invention has been described with reference to
exemplary embodiments, it will be understood by those skilled in
the art that various changes may be made and equivalents may be
substituted for elements thereof without departing from the scope
of the invention. In addition, many modifications may be made to
adapt a particular situation or material to the teachings of the
invention without departing from the essential scope thereof.
Therefore, it is intended that the invention not be limited to the
particular embodiments disclosed, but that the invention will
include all embodiments falling within the scope of the appended
claims. For example, blades 12 and disc 14 need not be configured
as specifically illustrated so long as they are part of a radial
compressor that benefits from tuning as described.
* * * * *